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April 14, 1964 s. R. TYLER v 3,128,822

HYDRAULIC SUPPLY SYSTEMS Filed Jan. 4, 1961 4.sheetssheet 1 ,s fg J6 z'il: la Y /90 l SCHEDULING THROTTLE BOOST A mil coNTRol. VALVE PUMP T,8./9 23 2/ |2/a 122 o F/G /lG s lill J5 F/G. 4

SOHEDULING CONTROL NOZZLE INVENTO TTORNEY.

April 14, 1964 s. R. TYLER 3,128,822

HYDRAULIC SUPPLY SYSTEMS Filed Jan. 4, 1961 4 Sheets-Sheet 2 INveNToR@MIMI wrm A-r-roRNEY S April 14, 1964 S. R. TYLER HYDRAULIC SUPPLYSYSTEMS Filed Jan. 4, 1961 4 Sheets-Sheet 5 SIMPLEX NOZZLE ATTORNEYApril 14, 1964 s, R. TYLER HYDRAULIC SUPPLY SYSTEMS Filed Jan. 4, 1961 4Sheets-Sheet 4 INVENTOR. A/VLEY I?. 7745/? A fr0/@V6 V5' Sha BY (n, l f"'1 Y l United States Patent O 3,l28,822 HYDRAULEC SUPPLY SYSTEMSStanley R. Tyler, 4d Jitnxrnside Road, Cheltenham, England Filed Jan.li, 1961, Ser. No. 80,967 14 Claims. (Cl. 15S-36.3)

This invention relates to variable pressure liquid supply systems andwhile it applies generally to pressure liquid systems it has particularapplication in connection with fuel supply systems for burners of thespray or vaporizing types as used in gas turbines, in afterburners, orin furnaces or the like for supplying atomized fuel at an accuratelycontrolled rate. Reference will be made hereinafter to the fuel supplyto the spill burners of a gas turbine engine, but it is to be understoodthat this is merely by way of example, and no limitation to such a usenor to this specific type of burner is to be implied. This applicationis a continuation-in-part of my earlier but copending application SerialNo. 676,637, filed August 6, 1957, and now abandoned.

The fuels employed for gas turbines need to be pumped at comparativelyhigh pressure to the spray burners and for this purpose positivedisplacement pumps are usually used. Since these fuels have very poorlubricating properties difficulties can arise in these pumps such asmechanical seizure or wear if the fuel is dirt contaminated. Thecentrifugal pump is the most desirable form of pump for supplying fuelunder pressure to spray burners, since the parts of the pump to whichthe fuel has access do not include any rubbing surfaces. However, underrestricted output conditions fuel in a centrifugal pump becomes heatedunduly due to the mechanical energy dissipated in it and one object ofthe present invention is to provide a fuel supply system using acentrifugal pump arranged to operate with a vapor core, and in whichfuel in the pump is not unnecessarily heated under restricted outputconditions. This arrangement is particularly useful for supplying reheatburners or afterburners in a gas turbine engine where the pressure fallssubstantially with reduced fuel flow. The improved eciency of a vaporcore pump at low pressures is then better realized. Also the pump can berun entirely empty of fuel thus eliminating heat rise in the fuel duringnon-reheat operation.

Spill spray nozzles for burners such as those mentioned above are incommon use which operate by supplying liquid fuel tangentially into aswirl chamber and from which that part of the fuel which is to be burnedescapes as a spray and the remainder returns to a lower pressure regionthrough spill orifices. These nozzles have the advantage over otherspray nozzles of giving adequate atomization over a very large range ofspray flow rates. However, when it is necessary to control accuratelythe rate of spray flow over a very Wide flow range with spill nozzles,it has heretofore been considered necessary to employ two pumps, one ofwhich (the supply pump) delivers metered fuel to a closed circuitoperating at high pressure, which includes the burner and the other pumpwhich is normally known as the circulating pump. The rst or supply pumpsupplies fuel in the quantity delivered at the burner nozzle. Thecirculating pump is difficult to construct for successful operationsince both inlet and outlet operate at high pressures and the mechanicaldrive to the pump must have a rotary seal operating at either inlet oroutlet pressure, being in either instance a high pressure rotary seal.Alternatively, leakage must be tolerated in order to reduce the pressureto a level acceptable to the seal. A further object of the presentinvention is to provide a centrifugal pump for use with spill spraynozzles which combines both supply and circulating pump features intoone pump and in which the high 3,128,822 Patented Apr. ld, 1964 ICCpressure rotary seal problem mentioned above does not exist,notwithstanding that a simple low pressure seal is required.

A still further object of the invention is to provide a simplecentrifugal Variable pressure liquid supply pump operable from any primemover or other power source. In accordance with the broad aspect of thepresent invention a variable pressure liquid supply system comprises acentrifugal pump suitably driven by a prime mover or the like and liquidflow control means acting to control the rate of flow of liquid enteringthe pump whereby there is insufficient liquid in the pump to iill it,and the contained liquid forms an annulus around the pump rotor, leavinga central hollow core, so that the output pressure is dependent on theradial depth of the annulus and the output rate of flow always tends tobe the same as the controlled input ow.

Another object of the invention is to provide a fuel supply system ofthe character described (vapor core) which will enable a wide range offlow rates to the burner to be attained accurately, without instability.

In accordance with a further aspect of the present invention a fuelsupply system for burner nozzles comprises a centrifugal pump to supplyliquid fuel to one or more spray nozzles combined with liquid fuel flowcontrol means controlling entry of liquid fuel into the centrifugal pumpso that a hollow core in the liquid fuel exists around the centre of thepump rotor whereby the generated pressure in the circulating systemincluding the spray nozzles is dependent on the radial depth which theliquid assumes around the rotor, this radial depth varying to generatepressure at the pump outlet to cause liquid fuel which enters the pumpto be pressurized to the extent that it will tend to leave the pump atthe rate at which it enters. It will be seen that if, for example, ahigher resistance to flow occurs in the spray nozzle then temporarilythe flow rate to the burner will drop allowing the radial depth of fuelin the pump rotor to increase thus increasing the pressure to overcomethe nozzle resistance so that the rate of flow to the nozzle is that ofliquid entering the pump. The converse also applies. In arranging acentrifugal pump to operate in this manner it will either be driven morequickly than a normal centrifugal pump to supply such a spray nozzle or,alternatively, it will be of a larger diameter. Where a centrifugal pumpis designed to operate full of liquid at a maximum flow rate andpressure, it is possible without alteration of the size of the pump orits driving speed to control the pump in accordance with the inventionto obtain reduced flow rates. Further, in accordance with the invention,where the spray nozzle is of the spill type (although the invention isalso applicable to simplex type burner nozzles and other fixed orificedevices) the high pressure supply to the nozzle is taken from the normalpump output at the circumference of the rotor via a diffuser while thespill return flow from the nozzle is fed through an annulus in the pumpcasing to a point upstream of the pump outlet, that is to say, radiallyinwardly of the outlet, at which point the pressure is less than thepressure at said outlet. Since the pressure in the liquid inthe pumpincreases with increase in radial depth it will be seen that the spillflow is thereby allowed to enter the pump rotor at a lower pressure zonefrom where it will move outwardly of the rotor and be pressurized to thehighest output pressure, without disturbing the hollow core. Thus,metered fuel fed to the inlet of such a centrifugal pump will all owfrom the spill spray nozzle as spray at the rate at which it enters thepump. Where the pump and nozzle are used in conjunction with a gasturbine engine and the pump is driven directly by the engine, it will beseen that the pressure difference in the centrifugal pump between theposition of entry of the spill flow and the pump outlet is proportionalto the square of the rotational speed. Also it is well-known that thedifference between inlet and spill pressures of a spill spray nozzle canbe made approximately proportional to the square of inlet ilow to thenozzle and from this it is clear that the nozzle inlet flow can be madeto vary substantially proportionally to the rotational speed of theengine. In this way the centrifugal pump in accordance with theinvention operates in a manner similar to that of a fixed displacementsupply pump feeding into a separate circulating system as normally usedwith spill spray nozzles.

In order that the invention may be clearly understood variousembodiments thereof will be described with reference to the accompanyingdrawings.

FIGURE 1 is a diagrammatic cross-section of a centrifugal pump andassociated control elements for feeding spill burners for use with a gasturbine engine.

FIGURES 2 and 3 are respectively cross-section and front elevationalviews of the centrifugal pump shown in FIGURE 1.

FIGURE 4 is a digrammatic view of an alternative centrifugal pumparrangement for use with spill burners.

FIGURES 5 and 6 are two further embodiments of centrifugal pumparrangements for use with ordinary burners of the simplex type (whereinall fuel permitted to reach the burner is discharged therefrom andburned) in a gas turbine engine.

FIGURE 7 is an axial sectional View of a practical form of centrifugalpump such as might be used, according to FIGURE 5, in supplying fuel toa simplex type burner.

Referring now to FIGURE 1, the centrifugal pump rotor is showndiagrammatically at 10 and is rotatably driven by means of a shaft 11.The rotor is mounted in the chamber 8 of a casing 12 and, being arrangedto supply a spill spray nozzle, the pump has adjacent to the rotor apair of circular channels 13 and 14 disposed in one wall adjacent to theopen sides of the rotor blades. Channel 13 is disposed adjacent to theperiphery of the rotor and is connected to an output pipe 15 leading tothe input connection 16 of a spill spray nozzle 17. An adjustablethrottle valve 15a, may be included in pipe 15. The spill flow returningfrom the nozzle 17 passes through ,pipe 18 and enters the channel 14 ofthe pump at a position slightly radially inwards of the channel 13.Thereby the channel 14 is of lower pressure than the channel 13. Theinlet 19 of the centrifugal pump is disposed at the centre thereof andfuel which enters the inlet 19 passes through a scheduling control 21and a boost pump 22 from a supply tank. A pressurizing valve 19a ininlet 19, which in effect is a spring-loaded non-return valve, is biasedto close against inflow of fuel from boost pump 22, hence insures thatno fuel will enter except so much as is required to make up the quantitydischarged at the nozzle 17, such quantity increasing the pressure dropacross the valve 19a. That valve also prevents vapor from passing backinto control 21. The boost pump delivers fuel from the tank at a lowpressure to the scheduling control and may comprise a small centrifugalpump or a fixed displacement pump having a valved relief passage (notshown, but conventional) extending from its delivery back to its inlet.The term scheduling control is used to indicate generally any of theusual controls, manual or automatic, which regulate fuel iloW inaccordance with operating conditions of the gas turbine. It will be inthe nature of a throttling device to govern the rate of fuel ow inaccordance with a parameter such as altitude, or in accordance with thedesired or constant speed of running of the engine, or in accordancewith increased or decreased rates of flow for acceleration ordeceleration. All such devices are known, hence its exact nature,construction, and mode of operation, and the nature of external partsthrough which it effects control, will necessarily vary widely inaccordance with the overall design of the system, and since the controlper se is not a part of this invention, it is shown onlydiagrammatically. In any event it senses the rate of fuel ilow andeffects control thereof or thereby, in accordance with operatingconditions. The scheduling control as shown in FIGURE 1 controls aliquid llow adjusting means 21a responsive to the scheduling control toadjust fuel flow by throttling. The scheduling control and liquid flowadjusting means deliver the fuel at the determined rate through pipe 23and past the pressurizing valve 19a to the inlet 19 of the pump. Thisvalve 19a insures a buildup of pressure in pipe 23 superior to pressureat inlet 19 before the valve will open to admit fuel to the pump. Thepump rotor is so designed and operated that if it were maintained fullof fuel it would be capable of delivering a slightly greater quantitythan the scheduling control is capable of delivering. As a result thepump is never completely filled with fuel during operation, since theliquid which enters the pump casing is centrifuged outwardly and formsan annulus around the casing, leaving a circular empty space or core atthe centre which is normally lled with fuel vapor at very low pressure.This hollow core is not vented nor connected to any other region. Thepressure in the channel 13 depends on the radial depth of the annulusand the speed of rotation of the pump. Delivered liquid is drawn offthrough pipe 15 and connection 16 to the spill burner 17. Spill liquidfrom the spill burner is delivered through pipe 18 to channel 14 whichis disposed radially inwards of the channel 13 in a zone of lowerpressure. It will be seen that spill liquid Ventering channel 14 isimmediately centrifuged outwardly and eventually again passes throughpipe 15 to the spill burner inlet. In fact a circulating systemoperating at high pressure is formed by spill burner 17, pipes 15 and 18and the outer part of the pump between channels 14 and 13. Liquid fed byscheduling control 21 into the pump, at the rate determined by thescheduling control, feeds into this circulating system by centrifugalaction of the rotor, but at relatively low pressure, with the resultthat the rate of entry of fuel into the pump is equal to the rate ofdischarge from the spill burner. If, during operation of the engine, theburner should suddenly increase resistance to ow of fuel due, forexample, to a partial blockage then the radial depth of fuel in the pumpwill build up until the pressure in pipe 15 is increased so that theflow is restored to the rate at which fuel is delivered to the pump.Thus fuel ow to the burners is substantially independent of conditionsat the burners. Leakage of fuel along the drive shaft 11 will not occurduring normal operation of the system, for while low pressure prevailsat the center of the pump, the centrifugal action urges all containedfuel outwardly, and strongly resists its inward movement. A low pressureseal 11a is provided on shaft 11 to prevent leakage either of air alongthe shaft into the pump, or of low presure liquid from the pump alongthe shaft. If blockage should increase the radial depth of the fuelannulus temporarily, even to the point where it reaches the shaft, thepressure here still remains low, hence any suitable and simple lowpressure seal 11a about the shaft, and not a high pressure seal, willsuffice. The throttle valve 15a may be used to restrict flow in pipe 15at high output flow rates from the pump in order to cause fuel to leavethe pump at channel 14 rather than to enter it so that the nozzle 17will eject as spray the fuel received from both pipes 15 and 18. The useof a throttle such as 15a to induce ow to a spill nozzle through itsspill pipe is disclosed in my copending application Serial No. 842,762,filed September 28, 1959, and now abandoned in favor of a continuationthereof led September 7, 1962, and assigned Serial No. 222,107.

Referring now to FIGURES 2 and 3, details are shown of the pumpappearing diagrammatically in FIGURE 1. It will be seen that the channel14 is quite close to the peripheral channel 13 of the pump. Pressure ineach is high, due to the centrifugal effect, although pressure inchannel 13 is the higher. A diffuser is shown in dotted lines by thedivergent fuel flow passage in delivery connection 24. Pressure isgained in the diffuser due to the speed reduction of fuel flowingthrough it. Connections to the channels 13 and 14 are shown in FIGURE 3at 24 and 25 respectively. The rotor herein shown is an entirelyconventional centrifugal rotor in which blades extend on one side of abackplate.

Referring now to FIGURE 4, an alternative arrangement for feeding aspill burner 26 is shown. Two centrifugal pumps having rotors 27 and 28are provided mounted in an extended chamber 6 of a single case 29 anddriven by a common shaft 31. The entry into the chamber 27 is by meansof a port 32 in the casing feeding into the centre of the rotor. Apressurizing valve 30 is located in port 32. Delivery from rotor 27 istaken from peripheral channel 33 around the rotor. Delivery from thischannel passes to passage 34 and into a scheduling control 35 of anyknown type. Fuel leaving the scheduling control passes through passage36 to the entry port 37 at the centre of rotor 28. Peripheral passage 38around rotor 28 receives liquid from rotor 28 and delivers it to pipe 39which leads to the inlet connection of spill burner 26. Spill flow fromspill burner 26 passes along the pipe 41 to the port 37. Fuel supplyinto the entry port 32 of rotor 27 is fed from a boost pump 42 whoseoutput is controlled by a servo control 43 which forms the liquid llowadjusting means. It may operate either on a by-pass valve where the pump42 is of fixed displacement or, alternatively, it may vary thedisplacement of pump 42 if it is a variable displacement pump. Pump 42draws liquid directly from the storage tank. A pipe 44 interconnects thescheduling control 35 with the servo control 43, the arrangement beingthat ow of fuel through the scheduling control 35 acts through pipe 44on the servo control to cause the required rate of ilow of fuel, ascalled for by the scheduling control, to be delivered by the pump 42.The pump rotor 28 is the circulating pump and forms a closed circulatingsystem with pipes 39 and 41 and the spill burner 26. Fuel under pressuredeveloped at the supply pump 27 and delivered by passage 34 is fed fromthe scheduling control 35 to this circulating system and eventuallyemerges from the spill nozzle as spray. The pump rotor 27 is madecomparatively large in diameter such that the rate of flow of fuel ofwhich boost pump 42 is capable does not fully lill the chamber of rotor27, the fuel forming an annulus around the rotor to leave a centralvapor core. Any leakage from the high pressure circulating pump rotor 28will only enter the chamber of rotor 27, at lower pressure. Any tendencyto leak along the shaft 31, outwardly of rotor 27, is readily counteredhy a low pressure seal 31a about this shaft. It will be seen that in thesame Way as described with reference to FIGURE l the pressure in channel33 around the rotor 27 will vary in accordance with the radial depth offuel and this pressure will vary automatically in order that the flowfrom channel 33 to pipe 34 is exactly in accordance with the scheduledfuel delivery into the rotor by boost pump 42. One essential differencefrom the arrangement of FIGURE l is that the scheduling control is nowplaced in the delivery line from the main centrifugal pump and acts tocontrol the delivery of the boost pump by liquid flow adjusting means43. The centrifugal pump rotor 28 acts in an entirely normal way and iscompletely primed with fuel.

It will be seen that in the arrangements of FIGURE l and FIGURE 4 thefact that the supply and spill lines to the spill burner, correspondingto the usual high pressure circulating system, are at high pressure doesnot necessitate the provision of rotary seals in the pump arrangements.A seal about the rotative shaft 11 or 31 normally is employed, but thisis a low pressure seal, and it is still correct to say that the highpressures prevalent in the supply and spill lines do not require thepresence of a rotary seal of the type necessary to contain such highpressures. There is no question of leakage of the 6 scheduled fuel flowby way of 15, 18 and burner 17 in the arrangement of FIGURE l, althoughit might take place to a limited extent in the arrangement of FIGURE 4.

In FIGURE 5 another system is diagrammatically shown for feeding into aspray nozzle 45 of the simplex type. In such a spray nozzle oil fueldelivered through the delivery pipe 57 passes immediately into a swirlcharnber and out of the spray orice. While reference has been made tothe nozzle 45 as being of the simplex type it may of course representother kinds of spray nozzle or fixed orifice device in which only onepipe connection extends to the nozzle carrying fuel under pressure. Thecentrifugal pump rotor 46 is driven by a shaft 47 from the engine and islocated in the chamber 4 of a casing 48 having a peripheral channel 49and a central inlet 5l. The inlet 51 is controlled by a valve 52 fromwhich a stem 53 extends to a servo piston and cylinder unit 54. Acompression spring 55 tends to hold the valve on its seat. Liquid fuelis supplied to the inlet 51 through a pipe 56 from a low pressure source(not shown). The pipe 57 extends from the channel 49 to the burner 45and in this pipe a scheduling control 58 is located. This schedulingcontrol comprises a cylinder '75 within which a waisted piston valvemember 76 is slidable under the inuence of a manually adjustable cam 77and a return spring 78. A port 79 within the cylinder wall is inunrestricted communication with the waisted portion of valve member 76and pipe 57 is connected to this port. A further port 81 is formedwithin the cylinder which is partially closed by the edge of land 82 ofthe valve member '75, adjustment of the valve member causing a throttleto be formed by port S1 which is a function of the setting of cam 77.The pipe 57 then extends from port 81 to the nozzle 45. Pipes 83 and S4extend from positions downstream and upstream of the throttle port S1and connect to either end of a cylinder 85 to cause movement of theincluded piston 86 against the load of a spring 87. A small diameterplunger 88 extends from piston 36 into bore 89 into which a pair ofports 91 and 92 open. A waisted portion 93 of plunger 8S controls fuelow between ports 91 and 92. Port 91 is connected through restrictor 94from pipe 84 Whilst port 92 is connected to low pressure through pipe95. This control operates on the servo piston and cylinder 54 andcontrol is exerted by means of a pipe 59 which extends from port 91.Fuel flowing through port 81 produces a pressure drop in accordance withow rate, this pressure drop acting through pipes 83 and 84 on piston 86in opposition to the load of spring S7. A small movement of piston 86will move plunger 88 from a completely closed to a completely openposition in its control of flow through ports 91 and 92 over whichmovement of the spring 87 is arranged to exert a substantially constantforce. Thus at a particular pressure drop at port 81 the plunger 38 willadjust flow through ports 91 and 92 and will thus determine a pressuredrop due to ow through restrictor 94. Pipe 59 connects the Variablepressure downstream of restrictor 94 to the servo piston and cylinderunit 54 and the opening given to valve 52 will result from thecompression of spring 55 by pressure from pipe 59. If the ow ratethrough pipe 57 is too large the excess pressure drop at port 81 causesmovement of piston 86 and plunger 88 to the left to permit greater owthrough ports 91 and 92 and thus to cause a greater pressure drop atrestrictor 94. This in turn will reduce pressure in pipe 59 causingspring 55 to move valve 52 to a more closed position to reduce entry offuel into the pump to compensate the excess output flow. A reducedoutput flow would cause opening of valve 52 to increase input flow.Since the cam 77 determinues the throttling effect of port 81 it will beclearly apparent that the setting of this cam will accurately determinethe entry of fuel ino the pump. Valve 52 will also perform the functionof a pressurizing valve included in the previous embodiments. As in theprevious embodiments, because of the excess delivery capacity of pumprotor 46 over the delivery capacity of pump 56, the liquid fuel in thepump forms an annulus leaving a central hollow core 4 and the adjustmentof radial depth of the annular liquid will cause pressure to bedeveloped at the pump outlet 57 appropriate to the operating conditionsso as to cause fuel to leave the pump at the rate at which it enters.

A furher alternative arrangement is shown in FIGURE 6, the object againbeing to supply fuel to a burner 63 of the simplex type. The centrifugalpump is arranged substantially as described with respect to FIGURE andincludes a rotor 64, and drive shaft 65, a casing 66, an output channel67 disposed peripherally around the casing and an inlet 68. In thisinstance a pipe 69 leading from channel 67 is connected directly to theburner 63. In the inlet opening 68 of the centrifugal pump aspringloaded valve 71 is provided which is spring-loaded to the closedposition, the loading being in a direction such that the fuel suppliedto the inlet through the pipe 72 must be at a higher pressure than fuelimmediately inside the pump inlet 68. Fuel ow is controlled by means ofa scheduling control 73 which exerts a servo control 102 on the deliveryof boost pump 74, this latter either having a servo control by-pass or aservo-controlled variable displacement mechanism. Fuel is delivered bymeans of the control 73 and pump 74 to pipe 72 at the rate determined byengine operating conditions to which the control 73 is sensitive.

In all of the described embodiments reference has only been made to onespray nozzle, but it will be appreciated that in most instances aplurality of spray nozzles are fed from one pump, the nozzles all beingparallel connected. Further, it will be seen that other types of burnermay equally well be supplied by apparatus in accordance with theinvention such for example as burners of the vaporizing type.

In the description above it has been assumed that the centrifugal pumpcasing is stationary and that the rotor therein rotates. Pumps are knownwherein the casing rotates and the internal member is non-rotative, butscoops up liquid rotated by the casing. These still function ascentrifugal pumps, and no restriction is intended herein in the broadaspects of the invention to one or the other style of centrifugal pump.

I claim as my invention:

1. In combination with a hydraulic load, a variable liquid supply systemcomprising drive means, relatively fixed and rotative elements arrangedcooperatively about an axis of rotation to form a centrifugal pumphaving a chamber and an inlet and an outlet therefor disposed adjacentthe axis and the periphery of the chamber, respectively, said pump beingconnected with the drive means to operate at a predetermined speed,means for supplying pressurized liquid to the pump, means defining inletand delivery passages for the pump interconnecting the liquid supplymeans and the hydraulic load with the chamber inlet and outlet,respectively, valve means disposed adjacent the chamber inlet andoperative to admit liquid to the chamber from the inlet passage inresponse to a pressure differential between such passage and thechamber, and control means operable to vary the rate of liquid flow inthe inlet passage in accord with a demand which is variable over apredetermined range of flow rates having an upper limit of less than thedelivery rate of the pump at said speed so that liquid flowthrough inthe pump chamber assumes the form of an annulus in the relativelyperipheral portion thereof, the relatively axial portion of the chamberbeing closed to atmosphere so that a hollow core is formed centrally ofthe liquid annulus to enable it to compensate for variation in thehydraulic load at a particular demand by adjusting its radial depth.

2. The combination according to claim l wherein the centrifugal pump hasa fixed housing defining its chamber and a rotor mounted in the chamberto perform as the rotative element.

3. The combination according to claim 1 further comprising schedulingcontrol means operable to sense the rate of liquid tiow in one of thepassages and to so regulate the rate of such flow in the inlet passage,through the medium of the first-mentioned control means, that the latterrate is maintained in accord with the demand.

4. The combination according to claim 3 wherein the scheduling controlmeans is operable to sense the rate of liquid flow in the inlet passage.

5. The combination according to claim 3 wherein the scheduling controlmeans is operable to sense the rate of liquid flow in the deliverypassage.

6. The combination according to claim 3 wherein the liquid supply meansincludes a boost pump disclosed in the inlet passage.

7. The combination according to claim 6 wherein the first-mentionedcontrol means includes a throttle valve disposed in the inlet passageintermediate the boost pump and the centrifugal pump chamber inlet. 8.The combination according to claim 6 wherein the first-mentioned controlmeans includes a servo control interconnecting the scheduling controlmeans with the boost pump and operable to regulate the delivery from thelatter pump.

9. The combination according to claim 3 wherein the valve means performsas the first-mentioned control means.

10. The combination according to claim 1 wherein the hydraulic loadincludes a spill nozzle having a chamber and an inlet, a dischargeoutlet, and a spill outlet therefor of which the inlet is interconnectedwith the pump chamber outlet by the delivery passage, said pump chamberalso having a second inlet disposed radially intermediate thefirst-mentioned pump chamber inlet and the pump chamber outlet, and saidpassage dening means further deiining a spill return passageinterconnecting the spill outlet of the spill nozzle chamber with saidsecond pump chamber inlet.

11. The combination according to claim l0 wherein the delivery passagehas a throttle valve therein.

l2. The combination according to claim l0 wherein the delivery passagehas diffuser means therein to effect a speed reduction in the liquidflow therethrough.

13. The combination according to claim l wherein the hydraulic loadincludes a simplex nozzle having a chamber and an inlet and a dischargeoutlet therefor of which the inlet is interconnected With the pumpchamber outlet by the delivery passage.

14. The combination according to claim l further comprising a secondcentrifugal pump in the delivery passage, said second centrifugal pumpalso being connected with the drive means to operate at said speed.

References Cited in the file of this patent UNITED STATES PATENTS787,039 Harris Apr. ll, 1905 1,353,915 Kime Sept. 28, 1920 2,547,959Miller Apr. 10, 1951 2,575,923 McMahan et al Nov. 20, 1951 2,658,330Carey Nov. 10, 1953 2,673,604 Lawrence Mar. 30, 1954 2,713,244. lChandler July 19, 1955 2,720,256 Pearson Oct. 1l, 1955 2,916,875 Morleyet al Dec. l5, 1959 FOREIGN PATENTS 1,184,654 France Feb. 9, 1959161,796 Great Britain Apr. 21, 1921

1. IN COMBINATION WITH A HYDRAULIC LOAD, A VARIABLE LIQUID SUPPLY SYSTEMCOMPRISING DRIVE MEANS, RELATIVELY FIXED AND ROTATIVE ELEMENTS ARRANGEDCOOPERATIVELY ABOUT AN AXIS OF ROTATION TO FORM A CENTRIFUGAL PUMPHAVING A CHAMBER AND AN INLET AND AN OUTLET THEREFOR DISPOSED ADJACENTTHE AXIS AND THE PERIPHERY OF THE CHAMBER, RESPECTIVELY, SAID PUMP BEINGCONNECTED WITH THE DRIVE MEANS TO OPERATE AT A PREDETERMINED SPEED,MEANS FOR SUPPLYING PRESSURIZED LIQUID TO THE PUMP, MEANS DEFINING INLETAND DELIVERY PASSAGES FOR THE PUMP INTERCONNECTING THE LIQUID SUPPLYMEANS AND THE HYDRAULIC LOAD WITH THE CHAMBER INLET AND OUTLET,RESPECTIVELY, VALVE MEANS DISPOSED ADJACENT THE CHAMBER INLET ANDOPERATIVE TO ADMIT LIQUID TO THE CHAMBER FROM THE INLET PASSAGE INRESPONSE TO A PRESSURE DIFFERENTIAL BETWEEN SUCH PASSAGE AND THECHAMBER, AND CONTROL MEANS OPERABLE TO VARY THE RATE OF LIQUID FLOW INTHE INLET PASSAGE IN ACCORD WITH A DEMAND WHICH IS VARIABLE OVER APREDETERMINED RANGE OF FLOW RATES HAVING AN UPPER LIMIT OF LESS THAN THEDELIVERY RATE OF THE PUMP AT SAID SPEED SO THAT LIQUID FLOWTHROUGH INTHE PUMP CHAMBER ASSUMES THE FORM OF AN ANNULUS IN THE RELATIVELYPERIPHERAL PORTION THEREOF, THE RELATIVELY AXIAL PORTION OF THE CHAMBERBEING CLOSED TO ATMOSPHERE SO THAT A HOLLOW CORE IS FORMED CENTRALLY OFTHE LIQUID ANNULUS TO ENABLE IT TO COMPENSATE FOR VARIATION IN THEHYDRAULIC LOAD AT A PARTICULAR DEMAND BY ADJUSTING ITS RADIAL DEPTH.